Method and apparatus for compressing a gas to a high pressure

ABSTRACT

A method is provided for compressing a gas in a single cycle and in a single cylinder to a pressure of at least 17.2 Mpa with a compression ratio of at least about five to one. The method further comprises dissipating heat from the cylinder during the compression stroke whereby the gas is discharged with a temperature significantly less than isentropic. The apparatus comprises a hollow cylinder and a reciprocable free-floating piston disposed therein. The piston divides the cylinder into: (a) a compression chamber within which a gas can be introduced, compressed, and discharged; and, (b) a drive chamber, into which a hydraulic fluid can be introduced and removed for actuating the piston. The apparatus further comprises a piston stroke length to piston diameter ratio of at least seven to one. For operating the apparatus with a compression ratio of at least five to one, an outlet pressure of at least 17.2 Mpa, and a gas discharge temperature significantly less than isentropic, the apparatus can further comprise a variable displacement hydraulic pump for controlling piston velocity, an electronic controller for maintaining an average piston velocity that is less than 0.5 feet per second, and a heat dissipator for dissipating heat from the cylinder.

FIELD OF THE INVENTION

The present invention relates to a method and apparatus for compressinga gas to a high pressure. More particularly, the method comprisescompressing a gas in a single cycle and in a single cylinder to a highpressure with a compression ratio of at least about five to one, whiledissipating heat from the cylinder during the compression stroke, anddischarging the gas with a temperature significantly less thanisentropic. The apparatus comprises a free-floating piston disposedwithin the cylinder and a piston stroke length to piston diameter of atleast seven to one.

BACKGROUND OF THE INVENTION

A conventional compressor that is operable to increase the pressure of agas to a high pressure by a ratio of more than four to one typicallyemploys two stages of compression. Conventional compressors operateunder near isentropic conditions and the use of multiple stages allowsheat exchangers, also known as intercoolers, to be employed betweenstages to cool the gas after each stage.

U.S. Pat. No. 5,863,186 (the '186 patent) discloses a method forcompressing gases using a multi-stage hydraulically driven compressor.The '186 patent discloses a method and apparatus that does not employintercoolers, but instead discloses a method of operating multiplecycles of each stage before the target output pressure for that stage isachieved. The '186 patent discloses using a cooling jacket to removeheat from the compressor. The compressor still compresses gas under nearisentropic conditions but the use of multiple cycles for each stageallows time for cooling the compressed gas, prior to the operation ofthe next stage. This arrangement does not allow continuous operation ofsuccessive stages because this would not allow sufficient time forcooling the gas between stages. Each stage begins after the previousstage is completed. In the preferred embodiment disclosed by the '186patent, two stages are employed to raise the pressure of the gas fromabout 150 to 500 psi (about 1.0 MPa to 3.4 MPa) to a compressor outputtarget pressure between 3000 to 6000 psi (about 20.7 MPa Mpa to 41.4MPa).

The cost of a multi-stage compressor increases with the number of stagesbecause separate compressor units are required for each stage. Eachcompression stage requires its own drive, piping and cooling stage,which adds to the manufacturing and maintenance costs associated withsuch multi-stage systems.

Conventional mechanically driven piston compressors that employ arotating crankshaft to drive the compressor piston are limited todesigns with relatively short piston strokes. Most mechanically drivenpiston compressors have cylinders with piston stroke length to pistondiameter ratios that are less than four to one, and more typically lessthan two to one. The piston stroke is defined herein as the distancetraveled by the piston between the beginning and end of the compressionstroke (that is, the maximum linear distance traveled by the piston inone direction). The piston diameter is essentially the same as thecylinder bore diameter. As used herein, the “length to diameter ratio”is defined as the ratio of the piston stroke length to the pistondiameter.

Mechanically driven piston compressors typically compensate for theirshort strokes by operating at high speeds, for example, in hundreds, ormore typically, in thousands of cycles per minute.

Known hydraulically driven reciprocating piston compressor systems thatemploy piston rods to connect the compressor pistons to a drive means,have also employed low length to diameter ratios (typically less thanfour to one). Low pressure compressors commonly employ a length todiameter ratio of about one to one. As the length to diameter ratioincreases, it becomes harder to maintain alignment of the piston rod andpiston, which can cause faster wearing around the seals. A higher lengthto diameter ratio also results in increased piston rod weight because ofthe increased rod length and the need to design against buckling. Acompressor cylinder with a higher length to diameter ratio also requiresa more elongated space to accommodate the compressor. That is, such acompressor requires an elongated space to accomodate an elongatedcylinder, the piston rod in the extended position, and an elongatedhydraulic cylinder. Notwithstanding the problems with alignment andweight, for some applications, such as the aforementioned vehicular fuelcompressor application, such an elongated space is not convenientlyavailable.

Free-floating piston compressors have been developed which use the samecylinder for a hydraulic drive chamber and a compression chamber.Free-floating piston compressors have no piston rod and the pistondivides the cylinder into the hydraulic drive chamber and thecompression chamber. During a compression stroke, hydraulic fluid isdirected to the drive chamber to actuate the piston and compress thefluid in the compression chamber. Conversely, during an intake stroke,hydraulic fluid exits from the drive chamber while fluid enters thecompression chamber. Some of the spatial limitations that are associatedwith employing a piston rod and external hydraulic drive can beaddressed by employing a free-floating piston, since the length of theapparatus is defined essentially by the length of the compressorcylinder, and the length of the apparatus is not compounded by thelength of an extended piston rod and a separate drive cylinder.Accordingly, a compressor that employs a free-floating piston can be atleast about half the length of a rod-driven compressor with the samebore and piston stroke.

However, high pressure gas compressors are unknown that can compress agas by a ratio of five to one or more, in a single cycle of a singlestage, and under significantly less than isentropic conditions. As thecompression ratio increases the cumulative temperature rise during thecompression cycle also increases, and under near isentropic conditionscompression is inefficient. For compressors with outlet pressuresgreater than about 2500 psi (17.2 MPa) and compression ratios greaterthan about four to one, compressors generally employ at least twostages, and some means for cooling the gas between stages.

An example of an application that requires a gas to be delivered at ahigh pressure is a fuel compressor system for an internal combustionengine. It is well known for natural gas fuelled engines to mix thegaseous fuel with intake air at relatively low pressures. However, morerecent developments have been made to inject gaseous fuel in dieselcycle engines, wherein the gaseous fuel is injected directly into thecombustion chamber late in the compression stroke. Compared to thepreviously described natural gas fuelled engines, these diesel cycleengines require the gaseous fuel to be compressed to a much higherpressure, for example, to overcome the in-cylinder pressure, to satisfymass flow requirements, and to improve mixing and penetration. By way ofexample, for such engines, a fuel compressor system receives fuel from asource, such as a storage tank or pipeline, and pressurizes the fuel toa pressure in the range of between about 3000 psi and about 3600 psi(between about 20.7 MPa and about 24.8 MPa), for direct injection intoan engine combustion chamber. Depending upon the available pressure fromthe fuel source, a single stage compressor operable to increase gaseousfuel pressure to injection pressure by a ratio of at least about five toone could replace the final two stages of a conventional multi-stagecompressor.

A compressor used for supplying fuel to an engine used as a prime moverfor a vehicle or for power generation has different design criteria thana compressor that is used for other applications, such as fillingstorage vessels. For example, for a vehicular application, a lightweight compressor apparatus can reduce vehicle weight and improveoverall vehicle efficiency, whereas, reduced weight can not have similarbenefits for a compressor installed at a stationary installation. Whilereliability, durability and efficiency are important for allapplications, these characteristics are of particular importance for acompressor used to supply fuel to an engine. A compressor failure canresult in costly downtime or stranding a vehicle, while inefficientoperation increases operating costs.

In addition, with an engine that is the prime mover for a vehicle,higher fuel consumption reduces vehicle range and limits the routes thata vehicle can be used for. Furthermore, range is also increased by usinga fuel compressor with a higher fuel compression ratio because thisincreases the amount of fuel that can be delivered from the fuel tank.

For engines used for power generation, the efficiency of each componenteffects overall efficiency, and low efficiency can have significanteconomic consequences when an engine is run on a continuous basis andunder high load conditions.

SUMMARY OF THE INVENTION

There is a need for a method of continuously compressing a gas to a highpressure by a ratio of at least about five to one in a single cycle of asingle compressor stage with a discharge gas temperature significantlyless than isentropic. The isentropic temperature is defined as thetheoretical temperature of a gas after compression when no heat isdissipated. A temperature significantly less than isentropic is definedherein as a gas temperature after compression that is higher than thegas temperature before compression but that is not high enough toinhibit the ability of the compressor to efficiently compress the gas tothe desired outlet pressure.

For example, a discharge temperature that is at least 25 degrees Celsiuslower than isentropic, and more preferably at least 50 degrees Celsiuslower than isentropic, would be considered a discharge temperature thatis significantly less than isentropic.

A method is provided of compressing a gas in a hydraulically drivenreciprocating piston compressor that comprises a cylinder, a freefloating piston disposed within the cylinder between a first closed endand a second closed end, a compression chamber defined by a volumewithin the cylinder between the first closed end and the piston, and adrive chamber defined by a volume within the cylinder between the secondclosed end and the piston. The method comprises:

-   -   (a) in an intake stroke,        -   supplying the gas to the compression chamber;        -   removing the hydraulic fluid from the drive chamber,        -   whereby the gas supplied to the compression chamber is at a            higher pressure than the hydraulic fluid within the drive            chamber causing the piston to move to reduce the volume of            the drive chamber and increase the volume of the compression            chamber until the compression chamber has expanded to a            desired volume and is filled with the gas; and    -   (b) in a compression stroke,        -   supplying the hydraulic fluid to the drive chamber whereby            the hydraulic fluid within the drive chamber is at a higher            pressure than the gas within the compression chamber,            causing the piston to move to increase the volume of the            drive chamber and reduce the volume of the compression            chamber thereby increasing the pressure of the gas held            within the compression chamber;        -   discharging the gas from the compression chamber when the            pressure of the gas is increased in a single cycle to a            pressure of at least 2500 psi (about 17.2 MPa), which is at            least about five times greater than the pressure of the gas            supplied to the compression chamber; and        -   dissipating heat from the cylinder during the compression            stroke whereby the gas is discharged from the compression            chamber with a temperature significantly less than            isentropic.

In a preferred method a piston stroke length to piston diameter ratio ofmore than seven to one is employed. Using a higher length to diameterratio provides more surface area for heat dissipation and shorter heatconduction paths within the cylinder chambers to the cylinder walls. Forexample, cylinders with a piston stroke length to piston diameter ratioof between ten to one and one hundred to one are possible. What issurprising with the preferred method is the amount of heat that can bedissipated during a compression stroke. Dissipating a significant amountof heat from the compressor cylinder allows gas to be compressed tohigher pressures and with higher compression ratios, compared toconventional compressors. As previously noted, for high pressure gascompression, compared to the present method, conventional methods employa plurality of compression stages with lower compression ratios andmeans for dissipating heat external to the compressor cylinder, forexample, with intercoolers, aftercoolers, and hydraulic fluid coolers.

A compressor cycle is defined by the completion of an intake stroke anda compression stroke. The speed of the compressor measured in cycles perminute also influences the ability of the apparatus to dissipate heatfrom the compression cylinder. Whereas the speed of conventionalcompressors has been generally governed by mass flow requirements (thatis the output capacity of the compressor), the present method operatingthe compressor at a speed that enhances heat dissipation. In thecompressor speed ranges within which conventional compressors operate,the speed does not have a significant effect on heat dissipation.According to the present method, compared to conventional compressors,when piston velocity and/or compressor speed (measured in cycles perminute) is reduced by about an order or magnitude, changes in pistonvelocity and compressor speed begin to have a significant effect on heatdissipation. According to the present method, compressor speed ispreferably no greater than 20 cycles per minute. For compressors withhigher length to diameter ratios, a compressor speed less than 20 cyclesper minute can result in a piston velocity of several feet per second,but as disclosed herein, the higher length to diameter ratio and lownumber of cycles per minute provide heat dissipation advantages thatoffset the disadvantages associated with a higher average pistonvelocity. Compressors with lower length to diameter ratios preferablyhave an average piston velocity less than 1.5 feet per second (about0.46 meters per second). For example, a compressor with a length todiameter ratio of about seven and a half to one preferably has anaverage piston velocity less than 0.5 feet per second (about 0.15 meterper second).

Operating at low speeds can also provide advantages resulting from lesscomponent wear and increased durability. Heat dissipation from thecylinder helps to keep piston ring seals at lower temperatures, whichcan be beneficial to reducing wear and degradation of materials.

Preferred embodiment of the method further comprises transferring heatfrom the cylinder to the ambient environment through a heat dissipator.An example of a heat dissipator is a cooling jacket disposed around thecylinder, wherein the method further comprises directing a coolant toflow through the cooling jacket. Heat dissipation is improved bymaintaining a coolant flow velocity through the cooling jacket thatensures there are no stagnant pockets within the cooling jacket. Highervelocities can also promote turbulence that enhances heat transfer tothe coolant from the cylinder wall. When the compressor is part of asystem that includes an engine, the coolant can be conveniently suppliedfrom the coolant reservoir of an engine coolant subsystem. However, thecoolant that circulates to an engine is generally too hot to have asubstantial effect as a coolant supplied to the compressor cylinder, andso when engine coolant is employed it preferably is supplied from acircuit that is independent from engine cooling circuits.

Instead of using a cooling jacket and a liquid coolant, the heatdissipator can comprise a plurality of thermally conductive finsprotruding from the cylinder. Such a heat dissipator operates byconducting heat from the cylinder to the plurality of fins, whichprovide a greater surface area for transferring heat to the ambientenvironment. When this type of heat dissipator is employed the methodcan further comprise blowing air through the plurality of fins toenhance heat dissipation.

The method of dissipating more heat from the compressor cylinder can becombined with controlling piston velocity during the compression stroke.In one embodiment of the method, the piston preferably travels with at afirst velocity during a first portion of the compression stroke and witha second velocity during a second portion of the compression stroke. Thesecond portion follows sequentially after the first portion and thesecond velocity is lower than the first velocity. Controlling pistonvelocity in this manner allows the piston to travel at a higher velocityduring the early part of the compression stroke when there is lesscumulative temperature rise, and at a slower velocity later in thecompression stroke when there is more cumulative temperature rise. Thetiming for changing from the first portion of the compression stroke tothe second portion of the compression stroke can be handled in a numberof ways. For example, this change can occur when an electroniccontroller determines that a predetermined criteria is satisfied suchas, for example, when gas pressure within the compression chamber or gasdischarge temperature exceeds a predetermined set point, or when thepiston is at a predetermined location within the cylinder.

Reducing piston velocity also helps to reduce component wear and methodsfor improving heat dissipation also reduce the operating temperature ofcomponents and seals, which can prolong their life (if such componentsdegrade over time with exposure to heat and/or thermal cycling).

The method can further comprise controlling piston velocity during adischarge portion of the compression stroke that occurs after the secondportion of the compression stroke. During the discharge portion of thecompression stroke, the gas pressure within the compression chamber isgreater than the gas pressure downstream from the compressor cylinderand gas is being discharged from the compression chamber. During thedischarge portion of the compression stroke, piston velocity ispreferably kept substantially constant. Piston velocity during thedischarge portion of the piston stroke is preferably equal to or lessthan piston velocity at the end of the second portion of the compressionstroke. Piston velocity can be controlled to follow a predeterminedspeed profile during the compression stroke. According to one method, aspeed profile can be selected from a plurality of predetermined speedprofiles to control piston velocity at different times during acompression stroke. Preferably the speed profile controls pistonvelocity to be highest near the beginning of the compression stroke withpiston velocity gradually declining to a lower velocity before stoppingat the end of the compression stroke. The differences between theplurality of predetermined speed profiles can be piston velocity atdifferent times and/or the rate that piston velocity changes during thecompression stroke. Of the plurality of predetermined speed profiles,the speed profile can be selected to maximize thermodynamic efficiencyof compression for the desired mass flow rate and compression ratio.

When gas is being discharged from the compressor, piston velocity can becontrolled to be substantially constant until near the end of the pistonstroke when piston velocity can be further reduced until the pistoneventually stops at the end of the compression stroke. According to thismethod, the power supplied to the hydraulic pump can fluctuate duringcompressor operation, depending upon how piston velocity is controlled.An objective of this method is controlling piston velocity to achieve adesired amount of heat dissipation.

The piston speed profile can be selected in response to a measuredoperating parameter. For example, the selected speed profile can beresponsive to desired mass flow rate, inlet gas pressure, desired gaspressure, and desired compression ratio.

A controller that operates the compressor can select a predeterminedspeed profile from a plurality of predetermined speed profiles. Of theavailable speed profiles, the selected speed profile preferablymaximizes thermodynamic efficiency of compression for the desired massflow rate and compression ratio.

In systems that operate for long periods of time in a steady statecondition, it is preferable for the power demands of system componentsto be substantially constant. Accordingly, with such systems, ratherthan controlling piston velocity, a preferred method further comprisessupplying a substantially constant amount of power to a hydraulic pumpduring a compression stroke. This can be achieved with a constant powerhydraulic pump. A consequence of operating in this manner is that pistonvelocity automatically decreases as gas pressure within the compressionchamber increases, which is beneficial for heat dissipation.

The present disclosure describes an apparatus for compressing a gas to ahigh pressure. The apparatus comprises a reciprocating piston compressorthat has a piston stroke length to piston diameter ratio of at leastseven to one. The apparatus is operable to compress a gas in a singlecycle of a single stage from a pressure of between about 300 to about600 psi (about 2.1 to about 4.1 MPa) to a pressure of between 2500 to5000 psi (about 17.2 to about 34.5 MPa) with a discharge gas temperaturesignificantly less than isentropic. Conventional compressors that areoperable to compress a gas to such high pressures typically do not havecompression ratios greater than about four to one. Compression ratioshigher than five to one are preferred because this allows a gas to becompressed to a high pressure using less stages. By way of example, withthe disclosed apparatus compression ratios between eight to one and tento one can be achieved. A number of features can be combined with theapparatus to facilitate its operation or to reduce discharge gastemperature further.

In particular, an apparatus for compressing a gas to a high pressurecomprises:

-   -   (a) a hollow cylinder;    -   (b) a free-floating piston reciprocable within the cylinder, the        piston dividing the cylinder into,        -   a compression chamber within which a gas can be introduced,            compressed, and discharged; and        -   a drive chamber, into which a hydraulic fluid can be            introduced and removed for actuating the piston; and    -   (c) a piston stoke length to piston diameter of at least seven        to one;        whereby the piston is operable to compress a gas by a ratio of        at least five to one in a single cycle to an outlet pressure of        at least 2500 psi (about 17.2 Mpa) with a discharge gas        temperature significantly less than isentropic.

The apparatus can further comprise a controller for maintaing an averagepiston velocity during a compression stroke that is less than 1.5 feetper second (0.46 meter per second). In some embodiments an averagepiston velocity of less than 0.5 feet per second (about 0.15 meter persecond) is preferred.

A variable displacement hydraulic pump can be employed for supplyinghydraulic fluid to the drive chamber. By changing hydraulic fluid flowrate piston velocity can be changed during a compression stroke. Theapparatus preferably further comprises a controller for controllinghydraulic pump displacement while operating the apparatus during acompression stroke. Accordingly in preferred embodiments, such acontroller is operable to control the hydraulic pump displacement toincrease, decrease or maintain the flowrate of hydraulic fluid into thedrive chamber, whereby piston velocity changes to predetermined speedsat predetermined times during a compression stroke. By way of example,the controller can be an electronic controller or a pre-calibratedmechanical controller. For example, in one embodiment, an electroniccontroller can be operable to control the hydraulic pump displacementwith response to measured parameters comprising at least one of gasdischarge temperature, gas pressure within the compression chamber, andpiston position within the compression cylinder.

Instead of a variable displacement hydraulic pump, a variable speedhydraulic pump can be employed, whereby piston velocity is controllableto increase or decrease piston velocity during a compression stroke. Forexample, piston velocity can be reduced by reducing the speed of thevariable speed hydraulic pump when gas pressure within the compressionchamber exceeds a predetermined set point.

In the alternative, as already disclosed with reference to the method,the apparatus can further comprise a constant power hydraulic pump forsupplying hydraulic fluid to the drive chamber.

A feature of the present invention is that it employs length to diameterratios that are higher than those typically employed by conventional gascompressors. Another advantage of a higher length to diameter ratio isthat it can facilitate reducing the proportion of dead space volume tototal cylinder volume, which helps to improve compressor efficiency.Preferably the dead space volume is less than 0.3% of total compressionchamber volume.

A higher length to diameter ratio also allows longer piston strokes andpotentially less cycles per minute for improved efficiency. A lowercompressor speed can be compensated for by a larger compression chambervolume, provided by an elongated cylinder. At lower compressor speeds,there are additional efficiency gains because there is less switching inthe hydraulic system, and with less cycles the dead space at the end ofthe piston compression stroke is not encountered as often.

An additional feature that can be combined with the apparatus is a heatdissipator for dissipating heat from the cylinder. The heat dissipatorsubstantially surrounds the cylinder for receiving and dissipating heatfrom the cylinder. In a preferred embodiment, the heat dissipatorcomprises a cooling jacket through which a coolant fluid can be directedto receive and remove heat therefrom. The cooling jacket preferablycomprises a shell structure spaced apart from the cylinder, and acoolant inlet associated with one end of the cylinder and a coolantoutlet associated with an opposite end of the cylinder, whereby coolantcan enter the cooling jacket through the coolant inlet and flow betweenthe shell and the cylinder to the coolant outlet.

In another preferred embodiment the heat dissipator comprises aplurality of fins protruding from the cylinder to conduct heat from thecylinder to the ambient environment. A fan can be added for directingair to flow between the plurality of fins to further increase heatdissipation.

The apparatus preferably comprises two cylinders that are operable intandem to supply a more continuous flow of high-pressure gas.

In a preferred embodiment of an apparatus comprising two cylinders, theapparatus comprises:

-   -   (a) a first reciprocating compressor comprising a first hollow        cylindrical body with fluidly sealed ends, a first free-floating        piston disposed within the first hollow cylindrical body        defining a first drive chamber having a hydraulic fluid port and        a first compression chamber having a gas port selectively        connectable to a low pressure gas supply system or a high        pressure gas system;    -   (b) a second reciprocating compressor comprising a second hollow        cylindrical body with fluidly sealed ends, a second        free-floating piston disposed within the second hollow        cylindrical body defining a second drive chamber and a second        compression chamber having a hydraulic fluid port and a first        compression chamber having a gas port selectively connectable to        the low pressure gas supply system or the high pressure gas        system;    -   (c) a hydraulic drive system that is operable to alternate        between:        -   supplying hydraulic fluid to the first drive chamber while            withdrawing hydraulic fluid from the second drive chamber;            and        -   withdrawing hydraulic fluid from the first drive chamber            while supplying hydraulic fluid to the second drive chamber;    -   whereby the first and second reciprocating compressors are        operable in tandem to increase the pressure of the gas by a        ratio of at least about five to one to a pressure of at least        about 2500 psi (about 17.2 MPa) with a discharge gas temperature        significantly less than isentropic.

The first and second reciprocating compressors preferably havesubstantially the same dimensions.

The hydraulic drive system can comprise a reversible hydraulic pump forreversing the direction of hydraulic fluid flow. In an alternativearrangement, the hydraulic drive system comprises a flow-switching valveoperable to selectively direct the hydraulic fluid to one of the firstand second drive chambers through the hydraulic fluid ports to cause acompression stroke while simultaneously receiving hydraulic fluid fromthe other one of the first and second drive chambers to cause an intakestroke.

As disclosed, the apparatus can be combined with one or more of thedisclosed features to reduce gas temperature and improve thermodynamicefficiency.

BRIEF DESCRIPTION OF THE DRAWINGS

The drawings illustrate specific embodiments of the invention but shouldnot be considered as restricting the spirit or scope of the invention inany way:

FIG. 1 is a schematic diagram of an apparatus for compressing gascomprising two hydraulically driven reciprocating compressors operatingin tandem.

FIG. 2 is a section view of a reciprocating compressor which illustratesa free-floating piston disposed in the compressor cylinder, with acooling jacket disposed around the compressor cylinder;

FIG. 3 is a section view of a reciprocating compressor which illustratesa free-floating piston disposed in the compressor cylinder with coolingfins extending radially from the compressor cylinder;

FIG. 4 shows an embodiment of a compressor that comprises a plurality ofhydraulically driven compression cylinders disposed within a commoncooling jacket;

FIG. 5 depicts graphs that show compression chamber pressure, pistonvelocity, and hydraulic pump power plotted against time whichcorresponds to piston travel during the course of one compressionstroke. The graphs represent an apparatus that employs a hydraulicsystem with constant power control.; and

FIG. 6 is a graph of experimental data that plots temperature riseagainst compressor speed. This graph shows that by using a compressorwith a length to diameter ratio of about 7.5:1, gas can be compressed tohigh pressures with a temperature gain that is significantly less thanisentropic if the piston velocity is reduced to a speed that allows timefor heat to be dissipated.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENT(S)

Referring to the drawings, FIG. 1 is a schematic diagram of a preferredapparatus for compressing gas comprising two hydraulically drivenreciprocating compressors 10 and 20. Compressors 10 and 20 operate intandem, with each compressor capable of increasing the pressure of afluid in a single cycle of each stage by a ratio of at least about fiveto one. For example, a gas can be compressed in such an apparatus froman inlet pressure of 500 or 600 psi (about 3.4 or about 4.1 MPa) to anoutlet pressure of at least 2500 to 3000 psi (about 17.2 to about 20.7MPa). Higher compression ratios are preferable because they allow thenumber of compression stages to be reduced. For example, compressionratios of between eight to one and ten to one are possible withembodiments of the disclosed apparatus to achieve outlet pressures ofbetween 2500 psi (about 17.2 Mpa) and 5000 psi (about 34.5 Mpa).

The embodiments of compressors described herein generally have length todiameter ratios of at least seven to one, but an equally importantfeature of these embodiments is that they are operable continuously witha discharge gas temperature significantly less than isentropic. By wayof example, the compressors schematically shown in FIG. 1 have a lengthto diameter ratio of about fifteen to one.

In the embodiment of FIG. 1, compressor 10 is made to the samespecifications as compressor 20, and they are substantially identical.In a compressor cycle of 360 degrees, the initiation of a compressionstroke in one of the compressors is offset from the initiation of acompression stroke in the other compressor by about 180 degrees. Thatis, the initiation of each piston stroke is roughly synchronized so thatwhen one compressor is beginning its compression stroke, the othercompressor is beginning its intake stroke. In practice, the pistoncompleting an intake stroke typically reaches the end of its strokeshortly before the other piston completes its compression stroke.

Free-floating piston 12 is movable within compressor cylinder 14 underthe influence of a pressure differential on opposite sides of piston 12.On one side of piston 12, cylinder 14 is filled with hydraulic fluid ina drive chamber and on the other side of piston 12, cylinder 14 isfilled with gas in a compression chamber. Cooling jacket 16 is spacedfrom cylinder 14, forming an annular cavity through which coolant canflow around cylinder 14 to dissipate heat therefrom. Sensor 18 isemployed to detect the position of piston 12.

Inlet pipe 30 is fluidly connected to compressor inlet ports fordirecting gas into the compressor compression chambers during respectiveintake strokes. One-way flow controllers 32 allow gas to enter therespective compression chambers from inlet pipe 30, and preventcompressed gas from flowing back into inlet pipe 30. The term “one-wayflow controller” when used herein will be understood by those skilled inthe art to be known types of flow controlling devices, generally knownas check valves, which permit fluid flow in one direction whilepreventing flow in the reverse direction, such as, for example, ballcheck valves, spring assisted ball check valves, wafer check valves,disc check valves, and compressor valves.

Discharge pipe 36 fluidly connects outlet ports from the compressorcompression chamber to a high-pressure system, such as, for example, afuel supply system for an engine. Such a fuel supply system can includean accumulator vessel that is filled with high-pressure gas to ensure asufficient supply is available. One-way flow controllers 38 allowcompressed gas to exit from the compressor chambers and flow todischarge pipe 36, while preventing gas delivered to discharge pipe fromreturning to the compressor chambers.

Coolant supply pipe 40 connects the cavity between cooling jacket 16 andcylinder 14 with a supply of coolant. Heat is transferable from cylinder14 to the coolant and warmed coolant is removed from the cavity throughan outlet connected to coolant return pipe 42, which returns the coolantto the cooling system. For example, when the compressor is used tosupply a high-pressure fuel to an engine, the cooling system employed tosupply coolant to the engine can also be employed to supply coolant tothe cooling jacket for the compressors. However, a separate cooling loopcan be employed if the engine coolant flowing in the engine coolant loopis not significantly cooler than the compressed gas. For example, anengine can have a separate cooling loop for turbocharger intercoolers,and the coolant flowing through such a loop can be significantly coolerthan the coolant that is used to cool the engine. An independent coolingloop is employed if the engine coolant is too hot. A higher temperaturedifferential between the coolant and the warmer compressed gas ispreferred, and in general, coolant is preferably supplied with atemperature less than 50 degrees Celsius.

The flow rate of the coolant is high enough to prevent local boiling ofthe coolant and to prevent stagnant pockets from forming within thecooling jacket cavity. Higher velocity flow also results in lesstemperature gain in the coolant, more turbulence in the boundary layernext to the cylinder wall, and higher heat transfer rates. Turbulenceincreases thermal conductivity from the cylinder to the coolant.

Hydraulic drive systems are well known but a preferred arrangement forthe compressor apparatus is a closed loop system. A closed loop designhelps to synchronize the movements of the pistons in the two compressorsand is also more efficient since the hydraulic fluid is delivered to thepump at high pressure from a drive chamber instead of at atmosphericpressure from a reservoir (as in the case of an open loop system).

Compressor operation is substantially the same for all embodiments.During the compression stroke, hydraulic fluid is directed to the drivechamber while a gas is compressed in the compression chamber. Ashydraulic fluid is introduced into the drive chamber, free-floatingpiston 12 advances within cylinder 14 to expand the volume of the drivechamber and reduce the volume of compression chamber.

In a preferred embodiment, the hydraulic pump is a horsepower limitedpump so the power required by the pump is substantially constant duringoperation and the velocity of piston 12 automatically changes during thecompression stroke so that it is fastest at the beginning of thecompression stroke and progressively slower until discharge pressure isreached. A lower piston velocity in the later part of the compressionstroke is advantageous for heat dissipation and achieving a dischargegas temperature significantly less than isentropic for efficientcompression of the gas. Generally, relatively little heat is generatedin the compression gas while the compression ratio remains below aboutthree to one, so during the early part of the compression stroke thepiston velocity can be higher since there is less need to provide timefor heat dissipation. Later in the compression stroke, when more heat isgenerated, a lower piston velocity allows more time for heatdissipation. This method of operation is discussed in greater detailbelow, with reference to FIG. 5.

In another preferred embodiment the hydraulic system employs a variabledisplacement hydraulic pump that can be controlled to change pistonvelocity for better heat dissipation. This method is also discussed ingreater detail below.

Position sensor 18 is used to determine when piston 12 is near the endof the compression stroke to signal when hydraulic fluid flow should bereversed. Position sensor 18 is preferably a sensor that can be mountedon the outside of the compressor body, for ease of maintenance and sothat the only ports required in the cylinder head are for fluid entryand exit. Many types of suitable sensors are known to persons skilled inthe art. For example, a magnetic switch can be employed to detect theposition of piston 12 near the end of the compression stroke.

Compressed gas exits cylinder 14 to discharge pipe 36 when the pressurewithin the compression chamber is greater than the pressure withindischarge pipe 36. In preferred embodiments, when the compressor isoperating at its maximum compression ratio, exit pressure of thecompressed gas is at least five times greater than the inlet pressure,and in some embodiments exit gas pressure can be between about seven andten times greater than the inlet pressure. One-way flow controllers suchas check valves 38, prevent pressurized gas from flowing back into thecompression chamber from discharge 36.

When the piston reaches the end of the compression stroke, the volume ofthe compression chamber at that point defines a “dead space”. The gasretained in the dead space is compressed to a high pressure but is notexpelled in the compression stroke.

Reciprocating piston compressors normally have a dead space, however,the larger the ratio of dead space to compression chamber volume, thelower the efficiency of the compressor. When the piston reversesdirection, the retained pressurized gas expands and fills the growingvolume of the compression chamber. For the initial portion of the intakestroke, the retained gas causes the pressure within the compressioncylinder to remain greater than the pressure in inlet pipe 30,preventing new gas from entering. A smaller dead space means more newgas can be drawn in from inlet pipe 30 during each intake stroke,resulting in higher compressor efficiency.

Compressors can be designed to reduce dead space by reducing the amountof cylinder length that corresponds to the dead space. If compressorcylinders of different lengths all have a dead space defined by acylinder length of between about ¹⁴ inch and 18 inch (about 6 to 3 mm),an advantage of a compressor with a higher length to diameter ratio isthat the cylinder length associated with the dead space represents asmaller fraction of compression chamber volumetric capacity. By way ofexample, if Compressor A has a cylinder length of 60 inches (1524 mm),and Compressor B has a cylinder length of 4 inches (102 mm), and bothcompressors have a dead space cylinder length of 1/8 inch (3 mm), inCompressor A, the dead space represents 0.2% of the cylinder volume,whereas in Compressor B, the dead space represents 3.1% of the cylindervolume. Accordingly, compressors with higher length to diameter ratioscan be more efficient because the dead space represents a smallerfraction of the compression chamber volumetric capacity for a givencylinder length of dead space. With the compressor configurationsdisclosed herein, a dead space volume that is less than or equal to 0.3%of the cylinder volume is preferred. For example, a cylinder with alength of 80 inches (about 2032 mm) and a dead space cylinder length of1/8 inch (about 3.2 mm) has a dead space volume that is 0.16% of thecylinder volume.

At the end of the compression stroke, piston 12 reverses direction. Totrigger the beginning of the intake stroke, the flow of hydraulic fluidis reversed, for example, by reversing the hydraulic fluid flow througha reversible pump, or by operating a flow switching device thatredirects the flow of hydraulic fluid between hydraulic fluid passagesso that the drive chamber that was connected to the hydraulic pumpdischarge during the compression stroke is now connected to the suctionof the hydraulic pump (in a closed loop system) or to a drain passage(in an open loop system). By design, during the intake stroke thepressure of the gas in inlet pipe 30 and the pressure of the gas withinthe dead space of the compression chamber is greater than the pressureof the hydraulic fluid in the drive chamber. As a result, piston 12moves under the influence of the gas pressure within the compressionchamber and piston 12 pushes the hydraulic fluid out of the drivechamber and into the suction of the hydraulic pump.

At the end of the intake stroke the compression chamber of cylinder 14is filled with gas from inlet pipe 30, and this gas is ready forcompression in the next compression stroke.

While the operation of compressor 10 alone has been described above, inthe preferred embodiment shown in FIG. 1, compressors 10 and 20 operatein tandem with their cycles offset by 180 degrees, so that whencompressor 10 is beginning its compression stroke, compressor 20 isbeginning its intake stroke, and vice versa. Pairing two compressors inthis manner allows a more continuous stream of compressed gas to besupplied to discharge pipe 36, in addition to providing a convenientarrangement for a closed loop hydraulic drive system.

FIG. 2 shows compressor 100, which illustrates a preferred embodiment ofa compressor with a length to diameter ratio of about eight to one.Compressor 100 comprises free-floating piston 112 disposed withincylinder 114, defining a compression chamber between piston 112 and endplate 120 and a drive chamber between piston 112 and end plate 122. Endplate 120 comprises bores 121 for respective inlet and outlet passagesfrom the compression chamber. One-way flow controllers can be installedwithin end plate 120 to control the direction of flow through passages121. End plate 122 comprises bore 123 through which hydraulic fluidflows into and out of the drive chamber. An advantage of the presentcompressor is its simplicity compared to conventional multi-stagecompressors. Hollow cylinders are easy to manufacture and are readilyavailable for purchase in specified lengths.

Free-floating piston 112 moves within cylinder 114 under the influenceof a pressure differential between the drive and compression chambers,as described with reference to FIG. 1. Ring seals 113 provide sealingbetween piston 112 and the interior surface of cylinder 114.

Free-floating piston 112 preferably reciprocates with an average cyclefrequency less than 20 cycles per minute. Higher cycle frequencies allowless time for cooling during compression. As disclosed above, pistonvelocity preferably changes during the compression stroke to enhanceheat dissipation. Compressor cycle frequency for a given flow capacityvaries according to the length to diameter ratio and the length of thepiston stroke. As will be discussed again with respect to the examplesset out below, in the course of a compression stroke a lower averagepiston velocity also allows more time for heat to be dissipated.However, as the length to diameter ratio increases, the compressor isbetter able to dissipate heat, and higher piston velocities can betolerated.

Cooling jacket 116 is spaced from and surrounds cylinder 114, providingan annular cavity through which a coolant can flow.

FIG. 3 shows compressor 200, which illustrates a preferred embodiment ofa compressor with a length to diameter ratio of about thirty to one. Byway of example, for a flow capacity of about 30 standard cubic feet perminute (about 0.8 standard cubic meters per minute), a compressorcylinder with a diameter of 1 inch (about 25.4 cm) and a length of about30 inches (762 cm) can be employed to raise the pressure of a gas froman inlet pressure of about 600 psi (about 4.1 MPa) to an outlet pressureof at least about 3000 psi (about 20.7 MPa).

Compressor 200 comprises free-floating piston 212 disposed withincylinder 214. Piston 212 defines a compression chamber between piston212 and end plate 220 and a drive chamber between piston 212 and endplate 222. Free-floating piston 212 moves within cylinder 214 under theinfluence of a pressure differential between the drive and compressionchambers, as described with reference to FIG. 1.

The heat dissipator in the embodiment of FIG. 3 comprises heatconductive fins 216 that radiate from cylinder 214. Heat is conductedaway from cylinder 214 and transferred from fins 216 to the coolerambient air. For applications that require enhanced cooling, air flowthrough fins 216 can be increased, for example, by using a fan (notshown) or by positioning cylinder 214 in a location where there is acool air flow.

Heat dissipation can be improved by employing smaller cylinderdiameters, which result in a shorter heat conduction path between thecenter of the cylinders and the cylinder walls. Higher length todiameter ratios also yield larger cylinder wall areas which results in alarger surface area for heat transfer. In compressor cylinders withhigher length to diameter ratios, these features combine to assist withheat dissipation, making compression with a discharge gas temperaturesignificantly less than isentropic possible. The following tableillustrates the effect of increasing the length to diameter ratio oncylinder wall area for a constant compression chamber volume. TABLE 1Relative Length to Cylinder Wall Diameter Ratio Area  1:1 1.00  4:1 1.59 8:1 2.00  9:1 2.08 10:1 2.15 15:1 2.47 30:1 3.11 50:1 3.68 100:1  4.64

As illustrated by Table 1, a length to diameter ratio equal to orgreater than eight to one results in at least twice as much surfacearea, compared to a cylinder with a piston stoke length to diameterratio of one to one. Since the amount of surface area continues toincrease as the length to diameter ratio increases, for improved heatdissipation, higher length to diameter ratios are preferred over lowerlength to diameter ratios.

Reciprocating piston compressors with very high length to diameterratios can be achieved by employing cylinders with smaller borediameters. For example, length to diameter ratios of between 50:1 and100:1 can be easily achieved with a bore diameter of ½ inch (about 13mm), and a length of between 25 inches (about 635 mm) for a 50:1 ratio,and 50 inches (about 1270 mm) for a 100:1 ratio. Such a small borediameter results in a relatively small cylinder volume so a plurality ofsmall bore cylinders can be combined to increase flow capacity.

FIG. 4 is an illustration of a plurality of compressor cylinders 400that are housed in common cooling jacket 410. Common gas distributionmanifolds (not shown) can be incorporated into an end plate that alsoseals an end of cooling jacket 410, or each cylinder can have its owninlet and outlet gas piping. An advantage of individual piping for eachcylinder is that the operation of each cylinder, or groups of cylinders,can be offset from one another to provide a more steady flow ofdischarge gas.

With cylinders that have smaller bore diameters if it is not be possibleto incorporate check valves into an individual end plate for eachcylinder, there will be a dead space volume associated with the pipingbetween the end of the cylinder bore and the check valve. However,because the internal diameter of such pipes is small relative to thevolume of the compression chamber, the dead space volume is alsorelatively small (compared to total cylinder volume).

The graphs of FIG. 5 illustrate a methods of controlling compressoroperation. In FIG. 5, the power drawn by the hydraulic pump issubstantially constant. While there are many ways to design a hydraulicsystem with substantially constant power requirements, one preferredexample is a system that employs a horsepower limited hydraulic pump.For example, when a compressor is employed to supply fuel to an engine,the engine typically provides the power needed to drive the hydraulicpump. That is, whether power to the pump is delivered mechanically (forexample, via a drive shaft or belts), or indirectly from electricalpower generated by the engine, which drives an electric motor, the powerused to operate the hydraulic pump is provided by the engine. When anengine is employed for power generation applications, engine stabilityand efficiency is improved by operating with less power fluctuations, soit is desirable to limit the maximum power of the hydraulic pump so thatit operates with substantially constant power requirements.

FIG. 5 shows the effect of using a horsepower limited hydraulic pump todrive a reciprocating piston compressor. The horizontal axis representstime with t1 being the beginning of the compression stroke and t3 beingthe end of the compression stroke.

Compression of the gas takes place between t1 and t2. The pressureincreases slowly at first and then more rapidly as the compressionstroke continues. Conversely, piston velocity is highest near thebeginning of the compression stroke, when gas pressure is lowest andthere is the least resistance to piston movement. Piston velocitydeclines as gas pressure increases.

Still with reference to FIG. 5, at t2, the discharge pressure isreached, and from t2 to t3 gas pressure is substantially constant as gasis discharged from the cylinder. Between t2 and t3 piston velocity isalso substantially constant, because constant gas pressure results inconstant resistance to piston movement.

In the embodiment of FIG. 5, throughout the compression stroke, thepower drawn by the hydraulic pump is substantially constant except atthe very beginning of the compression stroke where power requirementsmay be lower because of transient conditions.

A different method of operating the compressor comprises controllingpiston velocity to reduce gas discharge temperature to improve heatdissipation and thermodynamics of the compression process, whileaccepting higher fluctuations in power requirements. In this embodiment,gas compression occurs during two portions of the compression stroke.During the first portion of the compression stroke the objective is tomove the piston quickly since there is less temperature gain at lowcompression ratios. Accordingly, at the beginning of the compressionstroke, piston velocity is relatively high. The temperature of the gasis closer to isentropic because at higher piston velocities there isless time for heat to be dissipated, but this is tolerable because thecumulative temperature rise is relatively low. The power drawn by thehydraulic pump is at an intermediate level, because while the hydraulicfluid flow rate is high, the resistance is low since gas pressure islow.

In the second portion of the compression stroke gas pressure is elevatedto discharge pressure. In this portion of the compression stroke, thecumulative temperature rise begins to become more significant so pistonvelocity is reduced to allow more time for heat to dissipate. A balanceis selected between reducing piston velocity to achieve almostisothermal compression, and increasing compressor speed to achieve ahigher gas flow rate, while maintaining discharge gas temperaturesignificantly less than isentropic. During the second portion of thecompression stroke the power drawn by the hydraulic system increasesbecause when piston velocity is substantially constant, resistanceincreases as gas pressure increases.

During the last part of the compression stroke, the gas pressure equalsthe discharge pressure and gas is discharged from the cylinder as thepiston advances. The pressure during this part of the compression strokeis substantially constant A smooth discharge flow rate is preferred, sopiston velocity is preferably constant. Power requirements are alsosubstantially constant at constant pressure and substantially constantpiston velocity. The magnitude of the power requirement during thedischarge portion of the compression stroke depends upon thepredetermined discharge pressure (higher power requirements for higherdischarge pressures).

There are many ways, well known in the art for controlling hydraulicfluid flow rate and piston velocity. In one example, a variabledisplacement pump such as a swash plate pump with an adjustable swashplate angle can be employed.

In this embodiment, the power requirements for the gas compressor arenot constant. However, for some applications a variable compressor powerrequirement is not a problem. For example, when a gas compressor isemployed to supply fuel to an engine that is the prime mover for avehicle, because the load on a vehicle engine already variable, variablecompressor power requirements are also manageable. The compressor speedprofile during the compression stroke determines the efficiency of asystem that is operated according to this method. For example, the speedprofile for an individual compressor can be calibrated with regard togas intake pressure, gas discharge pressure, desired compression ratio,and mass flow requirements.

The timing for switching between the first portion of the compressionstroke and the second portion of the compression stroke can becontrolled in a number of ways.

In one embodiment, a flow meter measures the flow of hydraulic fluid tothe drive cylinders so that the position of the piston is known from theamount of hydraulic fluid that has been supplied. For example, when theflow meter measures an amount of hydraulic fluid that has a volume thatis equal to the volume of the drive chamber at the end of thecompression stroke, it is known that the piston is at the end of thecompression stroke. Such a flow meter can also be used to determinepiston position at intermediate points during the compression strokeallowing piston velocity to be controlled based upon piston position.

In other embodiments, other instruments can be employed to determinewhen piston velocity should be increased or decreased. By way ofexample, piston velocity can begin a compression stroke at apredetermined velocity, and a pressure sensor and/or temperature sensorcan be employed to determine when piston velocity should be decreased toallow more time for heat dissipation.

Those skilled in the art will understand that piston velocity can becontrolled to follow many speed profiles.

EXAMPLE 1

The graph shown in FIG. 6 represents data collected from a gascompressor that employed a free floating hydraulically driven piston.The compressor cylinder had a stroke length of 10 1/4 inches (about 261mm) and a bore diameter of 1 3/8 inches (about 34.9 mm), whichcorresponds to a length to diameter ratio of about 7.5:1. The cylinderwas cooled by ambient air that had a temperature of about 10 degreesCelsius.

The graph of FIG. 6 plots temperature rise in degrees Celsius on thevertical axis against compressor speed in cycles per minute. Nitrogengas was supplied to the compressor at a temperature of about 0 degreesCelsius.

Table 2 below sets out specific parameters associated with each of thedata points. TABLE 2 Compressor Speed (CPM) 18.8 14.4 9.4 4.8 InletPressure (MPa) 3.9 4.1 4.1 4.2 Outlet Pressure (MPa) 20.6 20.9 20.6 20.3Mass Flow (kg/hr) 25.8 19.6 12.6 6.7

Plotted as a straight line at about 160 degrees Celsius is thetemperature rise associated with isentropic conditions. The graphillustrates the following:

-   -   a) At compressor speeds lower than 20 cycles per minute, for the        same compressor operating with the same compression ratio, the        temperature rise measured in the discharged gas begins to        decrease as compressor speed decreases.    -   b) At compressor speeds higher than 20 cycles per minute, there        is no significant difference between the actual temperature rise        and the temperature rise that would be associated with        isentropic conditions. This shows that conventional piston        compressors, which operate at speeds much higher than 20 cycles        per minute, operate at near isentropic conditions, which limits        maximum compression ratios, and requires multiple compression        stages, intercoolers, and aftercoolers.

For the gas compressor of this example, a compressor speed of 20 cyclesper minute correlates to an average piston velocity of 0.57 feet persecond, and as shown by the graph of FIG. 6, piston velocity ispreferably still lower. For example, a compressor speed of about 5cycles per minute correlates to an average piston velocity of 0.14 feetper second. Conventional hydraulically driven piston compressors employpiston velocities that are orders of magnitude higher. At conventionalpiston velocities the benefits of reduced temperature rise in thecompression fluid is not realized, and there is no indication that suchbenefits can be significant until compressor speed is reduced well belowconventional levels.

EXAMPLE 2

The data set out in table 3 below was collected from three experimentsdone with a larger gas compressor that employed a free floatinghydraulically driven piston to compress natural gas. The compressorcylinder had a stroke length of 54 inches (about 1370 mm) and a borediameter of 2 1/2 inches (about 64 mm), which corresponds to a length todiameter ratio of about 21.6:1.

A coolant consisting of 50% glycol and 50% water was circulated througha cooling jacket surrounding the compressor cylinder. The temperature ofthe coolant supplied to the water jacket was about 15 degrees Celsius.

The hydraulic system employed a constant power hydraulic pump, resultingin piston velocity automatically decreasing as resistance to pistonmovement increased with increasing gas pressure.

The three experiments were done with different cycle frequencies(measured in cycles per minute) and different compression ratios. TABLE3 Experiment #1 #2 #3 Cycle Frequency 3 5 12 Average Piston 0.45 ft/s0.75 ft/s 1.6 ft/s Velocity (0.14 m/s) (0.23 m/s) (0.49 m/s) CompressionRatio 5.07 5.18  5.78 Gas Pressure (Inlet)  680 psig  690 psig  644 psigGas Pressure (Outlet) 3504 psig 3636 psig 3794 psig Gas Temperature 13.6° C.  13.6° C.  8.6° C. (Inlet) Gas Temperature 112.5° C. 126.3° C.137.9° C. (Outlet) Gas Temperature Rise  98.9° C. 112.8° C. 129.3° C.(actual) Temperature Rise 151.7° C. 153.9° C. 157.9° C. (Isentropic)Difference between  39.2° C.  27.6° C.  20.0° C. Actual Temperature Riseand Isentropic Temperature Rise

Even though the compression ratios are slightly different, the data intable 3 from experiments #1 and #2 illustrates that lower cyclefrequencies and a lower average piston velocity can be employed tosignificantly reduce the temperature rise during the compression stroke.In these experiments, a significant reduction in the temperature risewas achieved with an average piston velocity less than 0.75 feet persecond (about 0.23 meters per second). In experiment #3 some heat wasdissipated but gas discharge temperature was only 20 degrees Celsiusless than isentropic. Persons skilled in the art will understand thatadditional steps could be taken to reduce the discharge temperaturefurther. By way of example, reducing the temperature of the coolantsupplied to the water jacket or increasing the flow rate of the coolantare steps that can be taken to further reduce gas discharge temperature.

While reducing gas temperature rise has been disclosed as beingadvantageous for thermodynamic and energy efficiency, it is alsoimportant to note that reducing temperature rise also results in acooler apparatus, which is in itself beneficial. For example, theapparatus comprises moving parts that require dynamic seals. Theeffective life of dynamic seals is typically prolonged by maintainingthem at cooler temperatures during operation.

As will be apparent to those skilled in the art in the light of theforegoing disclosure, many alterations and modifications are possible inthe practice of this invention without departing from the spirit orscope thereof. Accordingly, the scope of the invention is to beconstrued in accordance with the substance defined by the followingclaims.

1. A method of compressing a gas in a hydraulically driven reciprocatingpiston compressor that comprises a cylinder; a free floating pistondisposed within said cylinder between a first closed end and a secondclosed end; a compression chamber defined by a volume within saidcylinder between said first closed end and said piston; and a drivechamber defined by a volume within said cylinder between said secondclosed end and said piston; said method comprising: (a) in an intakestroke, supplying said gas to said compression chamber; removing saidhydraulic fluid from said drive chamber, whereby said gas supplied tosaid compression chamber is at a higher pressure than said hydraulicfluid within said drive chamber, causing said piston to move to reducethe volume of said drive chamber and increase the volume of saidcompression chamber until said compression chamber has expanded to adesired volume and is filled with said gas; and (b) in a compressionstroke, supplying said hydraulic fluid to said drive chamber wherebysaid hydraulic fluid within said drive chamber is at a higher pressurethan said gas within said compression chamber, causing said piston tomove to increase the volume of said drive chamber and reduce the volumeof said compression chamber thereby increasing the pressure of said gasheld within said compression chamber; discharging said gas from saidcompression chamber when the pressure of said gas is increased in asingle cycle to a pressure of at least 2500 psi (17.2 MPa), which is atleast five times greater than the pressure of the gas supplied to saidcompression chamber; employing a piston stroke length to piston diameterratio of more than seven to one; and dissipating beat from said cylinderduring said compression stroke whereby said gas Is discharged from saidcompression chamber with a temperature significantly less thanisentropic.
 2. The method of claim 1 further comprising employing acylinder with a piston stroke length to piston diameter ratio of betweenten to one and one hundred to one.
 3. The method of claim 1 furthermaintaining an average piston velocity that is less than or equal to 1.5feet per second (0.46 meters per second).
 4. The method of claim 1further comprising transferring heat from said cylinder to said ambientenvironment through a heat dissipator.
 5. The method of claim 4 whereinsaid heat dissipator comprises a cooling jacket disposed around saidcylinder and directing a coolant to flow through said cooling jacket. 6.The method of claim 5 wherein coolant flows through said cooling jacketwith a velocity that ensures there are no stagnant pockets within saidcooling jacket.
 7. The method of claim 5 further comprising supplyingsaid gas to an engine and supplying said coolant from an engine coolantreservoir, but from a circuit that is independent from engine coolingcircuits.
 8. The method of claim 5 wherein said heat dissipatorcomprises a plurality of fins protruding from said cylinder and saidheat dissipator operates by conducting heat from said cylinder to saidplurality of fins which provides a greater surface area for transferringheat to the ambient environment.
 9. The method of claim 8 furthercomprising blowing air through said plurality of fins to increase heatdissipation.
 10. The method of claim 1 further comprising controllingwhen said piston reverses direction by sensing when said piston isproximate to an end of said cylinder.
 11. The method of claim 1 furthercomprising controlling piston velocity during said compression strokewhereby said piston travels with at a first velocity during a firstportion of the compression stroke and with a second velocity during asecond portion of the compression stroke, wherein said second portionfollows sequentially after said first portion and said second velocityis lower than said first velocity.
 12. The method of claim 11 furthercomprising changing from said first portion of said compression stroketo said second portion of said compression stroke when gas pressurewithin said compression chamber exceeds a predetermined set point. 13.The method of claim 11 further comprising controlling piston velocityduring a discharge portion of said compression stroke that occurs aftersaid second portion of said compression stroke when gas is beingdischarged from said compression chamber, wherein piston velocity duringsaid discharge portion is kept substantially constant.
 14. The method ofclaim 13 wherein said piston velocity during said discharge portion ofsaid compression stoke is equal to or less than piston velocity duringsaid second portion of said compression stroke.
 15. The method of claim1 further comprising controlling piston velocity to follow apredetermined speed profile.
 16. The method of claim 15 furthercomprising selecting said predetermined speed profile in response to ameasured operating parameter.
 17. The method of claim 16 wherein said inmeasured operating parameters include at least one of desired mass flowrate, inlet gas pressure, desired gas discharge pressure, and desiredcompression ratio.
 18. The method of claim 15 further comprisingselecting said predetermined speed profile from a plurality ofpredetermined speed profiles to control piston velocity at differenttimes during a compression stroke, wherein said speed profiles controlpiston velocity to be highest near the beginning of the said compressionstroke with piston velocity gradually declining to a lower velocitybefore stopping at the end of the compression stroke, the differencebetween said plurality of predetermined speed profiles can be the pistonvelocity at different times and/or the rate that piston velocity changesduring the compression stroke, wherein of said plurality ofpredetermined speed profiles, said selected predetermined speed profilemaximizes thermodynamic efficiency of compression for the desired massflow rate and compression ratio.
 19. The method of claim 1 furthercomprising gradually reducing piston velocity during a compressionstroke, until said gas is being discharged from said compressionchamber, and then maintaining a substantially constant piston velocityfor the remainder of said compression stroke.
 20. The method of claim 1further comprising supplying a substantially constant amount of power toa hydraulic pump during a compression stroke, whereby piston velocitydecreases as gas pressure within said compression chamber increases. 21.An apparatus for compressing a gas to a high pressure, said apparatuscomprising: (a) a hollow cylinder; (b) a free-floating pistonreciprocable within said cylinder, said piston dividing said cylinderinto: a compression chamber within which a gas can be introduced,compressed, and discharged; and a drive chamber, into which a hydraulicfluid can be introduced and removed for acing said piston; and (c) apiston stroke length to piston diameter of at least seven to one;whereby said piston is operable to compress a gas by a ratio of at leastfive to one in a single cycle to an outlet pressure of at least 2500 psi(17.2 Mpa) with a discharge gas temperature at least 25 degrees Celsiusless than isentropic.
 22. The apparatus of claim 21 further comprising acontroller for maintaing an average piston velocity during a compressionstroke that is less than 1.5 feet per second.
 23. The apparatus of claim21 wherein the ratio between piston stroke length and piston diameter isbetween ten to one and one hundred to one.
 24. The apparatus of claim 21further comprising a variable displacement hydraulic pump for supplyinghydraulic fluid to said drive chamber, whereby piston velocity ischangeable during a compression stroke.
 25. The apparatus of claim 24further comprising a controller for controlling hydraulic pumpdisplacement while operating said apparatus during a compression stroke.26. The apparatus of claim 25 wherein said controller is operable tocontrol said hydraulic pump displacement to increase, decrease ormaintain the flowrate of hydraulic fluid into said drive chamber,whereby piston velocity changes to predetermined speeds at predeterminedtimes during a compression stroke.
 27. The apparatus of claim 25 whereinsaid controller is operable to control the hydraulic pump displacementwith response to measured parameters comprising at least one of gasdischarge temperature, gas pressure within said compression chamber, andpiston position within said compression cylinder.
 28. The apparatus ofclaim 21 further comprising a constant power hydraulic pump forsupplying hydraulic fluid to said drive chamber.
 29. The apparatus ofclaim 21 wherein dead space volume is less than 0.3% of totalcompression chamber volume.
 30. The apparatus of claim 21 furthercomprising a heat dissipator for dissipating heat from the cylinder. 31.The apparatus of claim 30 wherein said heat dissipator substantiallysurrounds said cylinder for receiving and dissipating heat from saidcylinder.
 32. The apparatus of claim 31 wherein said heat dissipatorcomprises a cooling jacket through which a coolant fluid can be directedto receive and remove heat therefrom.
 33. The apparatus of claim 32wherein said cooling jacket comprises a shell structure spaced apartfrom said cylinder, and a coolant inlet associated with one end of saidcylinder and a coolant outlet associated with an opposite end of saidcylinder, whereby coolant can enter said cooling jacket through saidcoolant inlet and flow between said shell and said cylinder to saidcoolant outlet.
 34. The apparatus of claim 31 wherein said heatdissipator comprises a plurality of fins protruding from said cylinderto conduct heat from said cylinder to the ambient environment.
 35. Theapparatus of claim 34 wherein said heat dissipator further comprises afan for directing air to flow between said plurality of fins.
 36. Theapparatus of claim 21 wherein said apparatus comprises two cylindersthat are operable in tandem to supply a more continuous flow ofhigh-pressure gas.
 37. The apparatus of claim 21 further comprising agas inlet passage through which said gas is flowable into saidcompression chamber and a separate gas outlet passage through which saidgas is dischargeable from said compression chamber.
 38. The apparatus ofclaim 37 further comprising: a one-way flow controller for controllingthe one-way flow of said gas into said compression chamber through saidgas inlet passage; and a one-way flow controller for controlling theone-way flow of said gas out of said compression chamber through saidgas outlet passage.
 39. The apparatus of claim 38 wherein said gas inletand outlet passages pass through an end plate that seals saidcompression chamber and said one-way flow controllers are each disposedwithin said end plate.
 40. The apparatus of claim 21 wherein saidapparatus further comprises a sensor for detecting when said piston hascompleted a compression stroke.
 41. The apparatus of claim 21 whereinsaid apparatus is operable with a compression ratio of between eight toone and ten to one.
 42. A hydraulically driven reciprocatinghigh-pressure gas compressing apparatus comprises, (a) a firstreciprocating compressor comprising a first hollow cylindrical body withfluidly sealed ends, a first free-floating piston disposed within saidfirst hollow cylindrical body defining a first drive chamber having ahydraulic fluid port and a first compression chamber having a gas portselectively connectable to a low pressure gas supply system or a highpressure gas system; (b) a second reciprocating compressor comprising asecond hollow cylindrical body with fluidly sealed ends, a secondfree-floating piston disposed within said second hollow cylindrical bodydefining a second drive chamber and a second compression chamber havinga hydraulic fluid port and a first compression chamber having a gas portselectively connectable to said low pressure gas supply system or saidhigh pressure gas system; (c) a hydraulic drive system that is operableto alternate between: supplying hydraulic fluid to said first drivechamber while withdrawing hydraulic fluid from said second drivechamber; and removing hydraulic fluid from said first drive chamberwhile supplying hydraulic fluid to said second drive chamber; whereinfor each one of said hollow cylindrical bodies, the ratio between pistonstroke length and free-floating piston diameter is at least seven toone; whereby said first and second reciprocating compressors areoperable in tandem to increase the pressure of said gas by a ratio of atleast five to one to a pressure of at least 2500 psi (17.2 MPa) with adischarge gas temperature at least 25 degrees Celsius less thanisentropic.
 43. The apparatus of claim 42 wherein said hydraulic systemcomprises a reversible hydraulic pump for reversing the direction ofhydraulic fluid flow.
 44. The apparatus of claim 42 wherein saidhydraulic system comprises a flow switching valve operable toselectively direct said hydraulic fluid to one of said first aid seconddrive chambers through said hydraulic fluid ports to cause a compressionstroke while simultaneously receiving hydraulic fluid from the other oneof said first and second drive chambers to cause an intake stroke. 45.The apparatus of claim 42 wherein said first and second reciprocatingcompressors have substantially the same dimensions.
 46. The apparatus ofclaim 42 wherein said ratio between said piston stroke length and saidpiston diameter is between at least ten and up to and including onehundred to one.
 47. The apparatus of claim 42 further comprising a firstheat dissipator substantially surrounding said first cylindrical bodyand a second heat dissipator substantially surrounding said secondcylindrical body, whereby said first and second heat dissipators areoperable to transfer heat from each one of said cylindrical bodies to afluid which receives and removes heat from said compressor.
 48. Theapparatus of claim 47 wherein said first and second heat dissipatorscomprise respective cooling jackets through which a liquid coolant canflow to receive and remove heat from said compressors.
 49. The apparatusof claim 47 wherein said first and second heat dissipators each comprisea plurality of fins protruding from each one of said cylindrical bodiesto conduct heat from said cylinder to air in the ambient environment.50. The apparatus of claim 49 further comprising a fan for directing airto flow between said plurality of fins.
 51. The apparatus of claim 42wherein said gas entering said compressors is supplied with a pressureof between at least 300 psi (2.07 MPa) and up to and including 500 psi(3.45 Mpa).
 52. The apparatus of claim 42 further comprising sensors fordetecting when said free floating pistons reach respective end positionsand signaling a controller to reverse hydraulic fluid flow direction.53. The apparatus of claim 52 wherein said sensors employ a magneticswitch.
 54. The apparatus of claim 42 wherein said hydraulic systemcomprises a variable speed hydraulic pump whereby piston velocity iscontrollable to increase or decrease piston velocity during acompression stroke.
 55. The apparatus of claim 54 wherein pistonvelocity is reduced by reducing the speed of said variable speedhydraulic pump when gas pressure within the compression chamber exceedsa predetermined set point.
 56. The apparatus of claim 42 wherein saidhydraulic system comprises a constant power hydraulic pump.
 57. Anapparatus for compressing a gas to a high pressure, said apparatuscomprising: (a) a plurality of hollow cylinders; (b) a free-floatingpiston reciprocable within each one of said cylinders, said pistondividing each of said cylinders into. a compression chamber within whicha gas can be introduced, compressed, and discharged; and a drivechamber, into which a hydraulic fluid can be introduced and removed foractuating said piston; and (c) a piston stroke length to piston diameterof at least seven to one; and (d) a cooling jacket disposed around saidplurality of cylinders and comprising a fluid inlet and a fluid outletwhereby a coolant is flowable between said cylinders.
 58. The apparatusof claim 57 wherein each one of said cylinders can be employed tocompress a gas by a ratio of at least five to one in a single cycle toan outlet pressure of at least 2500 psi (17.2 Mpa) with a discharge gastemperature at least 25 degrees Celsius less than isentropic.
 59. Theapparatus of claim 57 wherein at least one of said pistons is operablewith a compression cycle that is offset from the other ones of saidplurality of pistons.